Turbochargers are unique mechanical devices in that they are expected to operate at extremely high RPM under conditions of high temperature and changing load, and yet are expected to provide long trouble-free service.
More specifically, a turbocharger is a type of forced induction system. Engine exhaust gases drive a turbine. The turbine is connected via a shaft to a compressor. Ambient air is compressed by the compressor and is fed into the intake manifold of the engine, allowing the engine to combust more fuel, and thus to produce more power for a given displacement. Considering the volumetric gas intake requirements of an engine operating at peak performance and the comparatively small size of a turbocharger, it can be appreciated that a turbocharger may be expected to rotate at speeds of up to 300,000 rpm.
The basic purpose of a bearing system is to provide a near frictionless environment to support and guide this rapidly rotating shaft over the life of the turbocharger, which should ideally correspond to that of the engine, which could be 500,000–1,000,000 km. The bearing system usually comprises two spaced-apart bearings, which function to dampen oscillations. Considering that the turbine is driven by engine exhaust gas, which may have a temperature as high as 1,300 F, it will be apparent that the bearing system must be designed so that a sufficient amount of lubricant is always channeled through the bearing system for removal of heat. Obviously, the turbocharger bearing system is a critical system that must be highly engineered.
On the other hand, it is highly desirable to design a turbocharger that is comprised of a minimum number of parts, which parts are easy to manufacture and easy to assemble, while still satisfying the demand for extended service life. Significant design effort has been directed toward improvements in turbocharger bearing systems.
In one popular turbocharger design the shaft is supported by a pair of floating radial bearings arranged in a cylindrical bore formed in the center housing (also referred to as the bearing housing) of the turbocharger. In this conventional bearing arrangement, the axial movement of each of the floating radial bearing is restricted by a pair of snap rings which are fitted into ring grooves formed on the inner wall of a cylindrical bore through the turbocharger center housing. See, for example, the turbocharger journal bearings described in U.S. Pat. Nos. 3,058,787 and 4,427,309. However, in the case wherein the floating radial bearings are axially restricted by a pair of snap rings, a problem occurs in that the end faces of the rapidly rotating floating radial bearings contact the stationary snap rings. This contact not only causes friction wear at the contact area, it may change the rotational speed of the bearing. In addition, a complicated machining process is necessary to form the four ring grooves on the inner wall of the cylindrical bore into which the snap rings must be seated, and, as a result, the manufacturing cost of the turbocharger is increased. Further, the seating of four snap rings is labor intensive. As the expected life of the engine increases, the turbocharger must be engineered for longer life.
An improvement came with the evolution of the “one piece” radial bearing assembly, in which the pair of floating radial bearings is connected by a cylindrical spacer. This eliminated the need for the respective inboard snap rings, and consequently reducing machining and assembly costs. Being one solid piece, this design was thought to provide good vibration damping characteristics. However, in such a radial bearing assembly, since the axial length of the radial bearing assembly is very long, and since the two bearings are rigidly connected and can not independently optimally adjust their position in the bore, there was a problem in that a complicated and precise machining process was necessary. In addition, since the bearing assembly is one continuous piece, any vibration due to shaft dynamics at one bearing end is instantly communicated to the other bearing end, and further, heat from the turbine side bearing is conducted through the thermally conductive metal spacer cylinder to the compressor side bearing. In addition, lubricating oil located in the sliding zone of the floating radial bearings cannot easily escape, and the friction loss of the floating radial bearings is increased.
In view of the above, U.S. Pat. No. 4,358,253 proposed to install a separate cylindrical “bearing spacer” axially between the pair of journal bearings. This bearing spacer was in the form of a tube in the space between a stationary housing and the rapidly rotating shaft. However, given the rapid flow of oil in this space, in order to stabilize and prevent “wobble” of the bearing spacer, the spacer was given an outer diameter corresponding substantially to that of the outer diameter of the bore. This greatly diminished or even completely stopped the rotation of the spacer, and thus prevented wobble. However, this spacer design tends to impede oil flow. Further, since the bearing spacer exhibits little or no rotational speed, wear is produced where the spacer contacts the rapidly rotating journal bearings. Further yet, given the high rotational speed of the shaft, the stationary spacer introduces drag and contributing to accelerated oil degradation in the space between the shaft and spacer.
It has also been proposed to utilize a bearing spacer having an inner diameter corresponding substantially to the outer diameter of the roatary shaft. While this snug fit would prevent wobble, such a close fit between bearing spacer and shaft causes the bearing spacer to rotate a high speed, causing shear and oil degradation, as well as drag on the shaft.
These prior designs utilizing a separate central bearing spacer have all proven satisfactory with regard to providing proper axial spacing of the radial bearings. However, the need to prevent wobble of the bearing spacer required the bearing spacer, if not integral with the bearings, to be either snugly fit to the shaft or snugly fit the bearing housing bore. These designs, though overcoming the problems associated with the four snap-ring design, have not provided adequate oil flow over and about the inner and outer diameter surfaces of the journal bearings and have not achieved satisfactory rotational speeds of the bearing spacer for reduction in drag, and as a result have suffered from relatively premature journal bearing failure.
As an improvement over the above described bearing spacers is provided in U.S. Pat. No. 4,902,144 entitled “Turbocharger Bearing Assembly”, teaching a bearing design employing a pair of journal bearings separated by a floating central spacer. The generally cylindrical, rotationally floating bearing spacer has opposite ends defining a pair of axially outwardly presented inboard thrust surfaces to maintain the two journal bearings in precision axial spaced relation. For radially locating or “piloting” the bearing spacer within the center housing bearing bore, the spacer exhibits pilot means radiating outwardly from the spacer outer diameter. This design allows unimpeded oil flow and thus achieves an improved oil flow over the journal bearings in comparison to the bearing system described in U.S. Pat. No. 4,358,253. However, the design of the bearing spacer is complex and thus is associated with manufacturing expense. Further, considering the changes in temperature, viscosity, and rotational speed of the turbocharger, it is difficult to design the spacer to have optimal rotational speeds over the entire rotational speed range of the turbocharger rotary shaft. Further yet, the three-piece design with the freely-floating bearing spacer lacks the inertia related stabilizing effect of the one-piece bearing spacer on any radial or rotational vibration of the journal bearings. Thus, one of the advantages of the “one piece” bearing system is missing in this “three piece” bearing system design.
Accordingly, there is a need for a simpler, easier to manufacture, lower cost bearing system for a turbocharger that achieves desired rotational dynamics of the three piece design, yet achieves the superior vibration damping characteristics of the one piece design, and yet does not suffer from the requirement for precise machining of the one-piece bearing.